Rotating fluid machine

ABSTRACT

A rotary valve of a rotating fluid machine is constructed by bringing, into contact on sliding faces, a moving valve plate and a stationary valve plate, and high pressure working medium passages and low pressure working medium passages penetrate mating faces of a valve body and the stationary valve plate. By sealing the outer circumference of the mating faces with a first sealing member and sealing the circumference of the low pressure working medium passages with a second sealing member inside the first sealing member in the radial direction, the pressure of a high pressure working medium leaking out of the high pressure working medium passages to the mating faces is caused to act on substantially the whole area of the mating faces, thereby preventing local deformation of the stationary valve plate and ensuring close contact between the sliding faces.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a rotating fluid machine provided witha casing, a rotor rotatably supported by the casing, a working sectionprovided on the rotor, and a rotary valve, provided between the casingand the rotor, for controlling the supply and discharge of the workingmedium to and from the working section.

2. Description of the Related Art

In a rotary valve for rotating fluid machines of this kind, a movingvalve plate provided on the rotor and a stationary valve plate providedon a valve body engaged with the casing to be unable to rotate andmovable in the direction of the axis of the rotor are brought intocontact with each other on sliding faces orthogonal to the axis, and therotation of the moving valve plate relative to the stationary valveplate causes steam of high temperature and high pressure to besuccessively supplied to or discharged from a group of axial pistoncylinders provided on the rotor.

The device described in Japanese Patent Laid-Open No. 2002-256805 hastwo sealing members on the mating faces of the valve body and thestationary valve plate so as to surround the high pressure workingmedium passages and low pressure working medium passages penetrating themating faces. These sealing members obstruct a high pressure workingmedium from leaking out of the high pressure working medium passages tobe discharged outside the mating faces or into the low pressure workingmedium passages.

In the above-described conventional device, with regard the mating facesof the valve body and the stationary valve plate, a high pressure actsin the radially inner direction on the annular sealing memberssurrounding the high pressure working medium passages, and a lowpressure acts in the radially inner direction on the annular sealingmembers surrounding the low pressure working medium passages. Therefore,there is a possibility that the high pressure and the low pressure actlocally on the stationary valve plate to invite flexural deformation,whereby the close contact between its sliding face and that of themoving valve plate is lost, so that high temperature high pressure steamleaks.

In view of the problems noted above, the present invention has an objectto eliminate the pressure imbalance on the mating faces of the valvebody and the stationary valve plate of the rotary valve in the rotatingfluid machine, thereby ensuring close contact between the sliding facesand the moving valve plate.

SUMMARY OF THE INVENTION

In order to achieve the object stated above, according to the presentinvention, there is proposed a rotating fluid machine comprising: acasing; a rotor rotatably supported by the casing; a working sectiondisposed on the rotor; and a rotary valve, provided between the casingand the rotor, for controlling the supply and discharge of a workingmedium to and from the working section, the rotary valve beingconstructed by bringing, into contact on sliding faces, a moving valveplate provided on the rotor and a stationary valve plate provided on avalve body, and high pressure working medium passages and low pressureworking medium passages penetrating mating faces of the valve body andthe stationary valve plate, wherein the outer circumference of themating faces is sealed with an annular first sealing member and thecircumference of the low pressure working medium passages is sealed withan annular second sealing member inside the first sealing member in theradial direction.

With the configuration described above, as the arrangement of theannular first sealing member on the outer circumference of the matingfaces of the valve body and the stationary valve plate of the rotaryvalve of the rotating fluid machine encloses both the high pressureworking medium passages and the low pressure working medium passagespenetrating the mating faces inside in their radial direction, thepressure of a high pressure working medium leaking out of the highpressure working medium passages to the mating faces is caused to act onsubstantially the whole area of the mating faces, thereby preventinglocal deformation of the stationary valve plate and ensuring closecontact between the sliding faces. Also, since the annular secondsealing member is arranged around the low pressure working mediumpassages inside the first sealing member in the radial direction, thesecond sealing member prevents the high pressure working medium on themating faces from being discharged to the low pressure working mediumpassages. As the second sealing member encloses only a narrow areaaround the low pressure working medium passages, there is no fear thatlow pressure locally deforms the stationary valve plate.

Further, according to a second feature of the present invention, inaddition to the first feature, the first and second sealing members areC-shaped seals each having an opening in its section, the opening of thefirst sealing member is arranged inward in the radial direction, and theopening of the second sealing member is arranged outward in the radialdirection.

With the configuration described above, as the opening of the firstsealing member comprising a C-shaped seal is arranged inward in theradial direction, the sealing performance of the first sealing membercan be increased by the pressure of a high pressure working mediumleaking from the high pressure working medium passages to the matingfaces, thereby preventing the high pressure working medium from beingdischarged outside from the mating faces. Further, as the opening of thesecond sealing member comprising a C-shaped seal is arranged outward inthe radial direction, the sealing performance of the second sealingmember can be increased by the pressure of a high pressure workingmedium leaking to the mating faces, thereby preventing the high pressureworking medium from being discharged out of the mating faces to the lowpressure working medium passages.

A group of axial piston cylinders 56 in an embodiment correspond to theworking section in the invention; sealing members 89 and 90 in theembodiment correspond to the first sealing member and the second sealingmember, respectively, in the invention; first and second steam passagesP1 and P2 in the embodiment correspond to the high pressure workingmedium passages in the invention; and the fifth and sixth steam passagesP5 and P6 in the embodiment correspond to the low pressure workingmedium passages in the invention.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a vertical sectional view of an expander according to apreferred embodiment.

FIG. 2 is a sectional view taken on line 2-2 in FIG. 1.

FIG. 3 is a view taken on line 3-3 in FIG. 1.

FIG. 4 is an enlarged view of Part 4 in FIG. 1.

FIG. 5 is an enlarged view of Part 5 in FIG. 1.

FIG. 6 is an exploded perspective view of a rotor.

FIG. 7 is a sectional view taken on line 7-7 in FIG. 4.

FIG. 8 is a sectional view taken on line 8-8 in FIG. 4.

FIG. 9 is an enlarged view of Part 9 in FIG. 4.

FIG. 10 is a sectional view taken on line 10-10 in FIG. 5.

FIG. 11 is a sectional view taken on line 11-11 in FIG. 5.

FIG. 12 is a sectional view taken on line 12-12 in FIG. 5.

FIG. 13 is a sectional view taken on line 13-13 in FIG. 5.

FIG. 14 is a view in arrowed direction 14 in FIG. 13.

FIG. 15 is a view in arrowed direction 15 in FIG. 13.

FIG. 16 is an enlarged view of Part 16 in FIG. 5.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

A preferred embodiment of the present invention will be described belowwith reference to the accompanying drawings.

As shown in FIG. 1 through FIG. 9, an expander E according to anembodiment is used in, for example, a Rankine cycle system. It convertsthermal energy and pressure energy of high-temperature high-pressuresteam as a working medium into mechanical energy and supplies theconverted energy. The casing 11 of the expander E is provided with acasing body 12, a front cover 15 connected to the front opening of thecasing body 12 with a plurality of bolts 14 . . . with a sealing member13 therebetween, a rear cover 18 fitted to the rear opening of thecasing body 12 with a plurality of bolts 17 . . . with a sealing member16 therebetween, and an oil pan 21 fitted to the bottom opening of thecasing body 12 with a plurality of bolts 20 . . . with a sealing member19 therebetween.

A rotor 22 is arranged to be rotatable around an axis L extending in themiddle of the casing 11 in the back and forth directions, and supportedin front by combined angular bearings 23 f and 23 r disposed on thefront cover 15 and on the back by a radial bearing 24 disposed on thecasing body 12. A swash plate holder 28 is integrally formed on the rearface of the front cover 15. A swash plate 31 is rotatably supported bythis swash plate holder 28 via an angular bearing 30. The axis of theswash plate 31 is inclined relative to the axis L of the rotor 22 at afixed angle.

The rotor 22 is provided with an output shaft 32 supported on the frontcover 15 with the combined angular bearings 23 f and 23 r, three sleevesupporting flanges 33, 34 and 35 formed integrally with one another onthe rear part of the output shaft 32 via notches 57 and 58 of apredetermined width (see FIG. 4 and FIG. 9), a rotor head 38 connectedto the rear sleeve supporting flange 35 with a plurality of bolts 37 . .. via a metal gasket 36 integrally and supported on the casing body 12by the radial bearing 24, and a thermally insulating cover 40 fittedonto the three sleeve supporting flanges 33, 34 and 35 from front andconnected with a plurality of bolts 39 . . . to the front sleevesupporting flange 33.

Five sleeve supporting holes 33 a . . . , 34 a . . . and 35 a . . . arerespectively bored in the three sleeve supporting flanges 33, 34 and 35around the axis L at 72° intervals. Five cylinder sleeves 41 . . . arefitted into the respective sleeve supporting holes 33 a . . . , 34 a . .. and 35 a . . . from behind. Formed at the rear end of each of thecylinder sleeves 41 is a flange 41 a, which is positioned in the axialdirection in contact with the metal gasket 36 in a state in which it isfitted onto a stepped portion 35 b formed in the sleeve supporting hole35 a of the rear sleeve supporting flange 35 (see FIG. 9). A piston 42is slidably fitted within each of the cylinder sleeves 41, the front endof the piston 42 is in contact with a dimple 31 a formed in the swashplate 31, and a steam expansion chamber 43 is partitioned between therear end of the piston 42 and the rotor head 38.

A plate-shaped bearing holder 92 is laid over the front face of thefront cover 15 with a sealing member 91 therebetween and fixed withbolts 93 . . . . A pump body 95 is laid over the front face of thebearing holder 92 with a sealing member 94 therebetween and fixed withbolts 96 . . . . The combined angular bearings 23 f and 23 r arepositioned between the stepped portion of the front cover 15 and thebearing holder 92, and fixed in the direction of the axis L.

A shim 97 of a predetermined thickness is placed between a flange 32 dformed in the output shaft 32 supporting the combined angular bearings23 f and 23 r and the inner races of the combined angular bearings 23 fand 23 r. The inner races of the combined angular bearings 23 f and 23 rare fastened with nuts 98 screwed onto the outer circumference of theoutput shaft 32. As a result, the output shaft 32 is positioned in thedirection of the axis L relative to the combined angular bearings 23 fand 23 r, namely with respect to the casing 11.

The combined angular bearings 23 f and 23 r are attached in mutuallyreverse orientations, and support the output shaft 32 not only in theradial direction but also immovably in the direction of the axis L.Thus, one combined angular bearing 23 f is arranged to restrict theforward movement of the output shaft 32, while the other combinedangular bearing 23 r is arranged to restrict the backward movement ofthe output shaft 32.

As the combined angular bearings 23 f and 23 r are used as a bearing forthe front part of the rotor 22, one of the loads arising toward theopposite ends of the axis L in the expansion chambers 43 . . . in apredetermined operating state of the expander E is transmitted via therotor 22 to the inner races of the combined angular bearings 23 f and 23r, and the other load is transmitted via the swash plate 31 and theswash plate holder 28 of the front cover 15 to the outer races of thecombined angular bearings 23 f and 23 r. These two loads compress theswash plate holder 28 of the front cover 15 held between the angularbearing 30 supporting the swash plate 31 and the combined angularbearings 23 f and 23 r supporting the rotor 22, resulting in an enhancedrigidity of the mechanism. Moreover, the integral configuration of theswash plate holder 28 with the front cover 15 as in this embodiment ofthe invention makes the structure more rigid and simpler.

Further, by incorporating the angular bearing 30 supporting the swashplate 31 and the combined angular bearings 23 f and 23 r supporting therotor 22 into the front cover 15, it is possible to accomplish theassembling process in the units of “the rotor 22 and the piston 42 . . .”, “assembly of the front cover 15” and “the pump body 95” therebyimproving the efficiency of procedures such as rearrangement of thepiston 42 . . . and the replacement of an oil pump 49.

The radial bearing 24 supporting the rotor head 38 which constitutes therear end of the rotor 22 is an ordinary ball bearing supporting only theload in the radial direction. To enable the rotor head 38 to slide inthe direction of the axis L relative to the radial bearing 24, a gap ais formed between the rotor head 38 and the inner race of the radialbearing 24 (see FIG. 5).

An oil passage 32 a extending on the axis L is formed within the outputshaft 32 integral with the rotor 22. The front end of the oil passage 32a branches in radial directions to communicate with an annular groove 32b on the outer circumference of the output shaft 32. In a radially innerposition of the sleeve supporting flange 34 at the center of the rotor22, an oil passage blocking member 45 is screwed into the innercircumference of the oil passage 32 a with a sealing member 44therebetween. A plurality of oil holes 32 c . . . extend from the nearbyoil passage 32 a outward in the radial direction, and open in the outercircumferential face of the output shaft 32.

A trochoidal oil pump 49 is arranged between a concave 95 a formed inthe front face of the pump body 95 and a pump cover 48 fixed with aplurality of bolts 47 . . . to the front face of the pump body 95 with asealing member 46 therebetween, and includes an outer rotor 50 rotatablyfitted into the concave 95 a, and an inner rotor 51 fixed to the outercircumference of the output shaft 32 to engage with the outer rotor 50.The inner space of the oil pan 21 communicates with the intake port 53of the oil pump 49 via an oil pipe 52 and the oil passage 95 b of thepump body 95. The discharge port 54 of the oil pump 49 communicates withthe annular groove 32 b of the output shaft 32 via the oil passage 95 cof the pump body 95.

The piston 42 slidably fitted into the cylinder sleeve 41 consists of anend portion 61, a middle portion 62 and a top portion 63. The endportion 61 is a member having a spherical portion 61 a in contact withthe dimple 31 a of the swash plate 31, and is welded onto the tip of themiddle portion 62. The middle portion 62 is a cylindrical member havinga large-capacity hollow space 62 a, and has in the outer circumferentialpart near the top portion 63 a smaller diameter part 62 b slightlyreduced in diameter. A plurality of oil holes 62 c . . . are formed topenetrate the smaller diameter part 62 b in the radial direction. Aplurality of spiral oil grooves 62 d . . . are formed in the outercircumferential part ahead of the smaller diameter part 62 b. The topportion 63 facing the expansion chambers 43 is formed integrally withthe middle portion 62. A thermally insulating space 65 (see FIG. 9) isformed between a partition wall 63 a formed inside the space and a lidmember 64 fitted and welded onto its rear end face. Fitted to the outercircumference of the top portion 63 are two compression rings 66 and oneoil ring 67. An oil ring groove 63 b into which the oil ring 67 isfitted, communicates via a plurality of oil holes 63 c . . . with thehollow space 62 a of the middle portion 62.

The end portion 61 and the middle portion 62 of the piston 42 are madeof high carbon steel, and the top portion 63, of stainless steel. Theend portion 61 undergoes induction quenching, and the middle portion 62,plain quenching. As a result, the piston 42 obtains a high surfacestress resistance in the end portion 61 which is in contact with theswash plate 31 under a high surface stress, a wear resistance in themiddle portion 62 which is in sliding contact with the cylinder sleeves41 under poor lubricating conditions, and a heat and corrosionresistance in the top portion 63 which faces the expansion chambers 43to be exposed to high temperature and high pressure.

An annual groove 41 b (see FIG. 6 and FIG. 9) is formed in the outercircumference of the middle portion of each cylinder sleeve 41, and aplurality of oil holes 41 c . . . are formed in this annual groove 41 b.Irrespective of the mounting position of the cylinder sleeve 41 in therotating direction, the oil holes 32 c . . . formed in the output shaft32 and oil holes 34 b . . . (see FIG. 4 and FIG. 6) formed in the middlesleeve supporting flange 34 of the rotor 22 communicate with the annualgroove 41 b. A space 68 formed between the thermally insulating cover 40and the sleeve supporting flanges 33 and 35 respectively before andbehind the rotor 22 communicates with the inner space of the casing 11via oil holes 40 a . . . (see FIG. 4 and FIG. 7) formed in the thermallyinsulating cover 40.

An annular lid member 69 is welded onto the front side of the rotor head38 connected with the bolts 37 . . . to the rear face of the sleevesupporting flange 33 in the front side of the rotor 22, or onto theexpansion chambers 43 . . . An annular thermally insulating space 70(see FIG. 9) is defined on the back or rear face of the lid member 69.The rotor head 38 is positioned in the rotating direction by a knock pin55 relative to the rear sleeve supporting flange 35.

The five cylinder sleeves 41 . . . and the five pistons 42 . . .constitute a group of axial piston cylinders 56 according to the presentinvention.

Next will be described with reference to FIG. 5 and FIG. 10 through FIG.15 the structure of a rotary valve 71 for supplying and dischargingsteam to and from the five expansion chambers 43 . . . of the rotor 22.

As shown in FIG. 5, the rotary valve 71 arranged along the axis L of therotor 22 is provided with a valve body 72, a stationary valve plate 73,and a moving valve plate 74. The moving valve plate 74, in a state ofbeing positioned by a knock pin 75 in the rotating direction on the rearside of the rotor 22, is fixed with bolts 76 screwed onto the oilpassage blocking member 45 (see FIG. 4). The bolts 76 also have afunction to fix the rotor head 38 to the output shaft 32.

As is clear from FIG. 5, the stationary valve plate 73 in contact withthe moving valve plate 74 via the flat sliding faces 77 is fixed to thecenter of the front face of the valve body 72 with a single bolt 78, andfixed to the outer circumference of the valve body 72 with an annularfixed ring 79 and a plurality of bolts 80. When it is fixed, a steppedportion 79 a formed on the inner circumference of the fixed ring 79 ispressed onto the outer circumference of the stationary valve plate 73 ina spigot-fit manner, and a stepped portion 79 b formed on the outercircumference of the fixed ring 79 is spigot-fitted onto the outercircumference of the valve body 72, thereby ensuring a coaxialrelationship of the stationary valve plate 73 to the valve body 72.Further, a knock pin 81 for positioning the stationary valve plate 73 inthe rotational direction is arranged between the valve body 72 and thestationary valve plate 73.

Therefore, as the rotor 22 turns, the moving valve plate 74 and thestationary valve plate 73 turn relative to each other in close contactwith each other on the sliding faces 77. The stationary valve plate 73and the moving valve plate 74 are made of a highly durable material,such as carbon or ceramic, and their durability can be further enhancedby affixing a member having excellent heat resistance, lubricatingperformance, corrosion resistance and wear resistance to the slidingfaces 77, or by coating them with such a material.

The valve body 72 made of stainless steel is a stepped columnar memberhaving a larger diameter part 72 a and a smaller diameter part 72 b. Theouter circumferential faces of those larger diameter part 72 a andsmaller diameter part 72 b are respectively fitted onto the supportingfaces 18 a and 18 b having a circular section in the rear cover 18 withsealing members 82 and 83 therebetween to be slidable in the directionof the axis L.

As is clear from FIG. 5 and FIG. 15 referenced together, threesupporting holes 18 c . . . are formed in the inner face of the rearcover 18 opposite a stepped portion 72 c of the valve body 72. Thelarger diameter parts 84 a . . . of three engaging pins 84 . . . arepressed into these supporting holes 18 c . . . while the smallerdiameter parts 84 b . . . of those engaging pins 84 . . . are slidablyfitted into three pin holes 72 d . . . formed in the stepped portion 72c of the valve body 72. The three engaging pins 84 . . . are arranged at120° intervals on the same circle so as to surround the axis L. Whilerestricting the valve body 72 in the rotational direction with theseengaging pins 84 . . . thereby preventing them from being dragged by andaccompanying the rotor 22 in rotation, the valve body 72 is made movablein the direction of the axis L to secure the close contact between thesliding faces 77.

Referring again to FIG. 5, a plurality of preload springs 85 aresupported by the rear cover 18 so as to surround the axis L, and thevalve body 72 in which the stepped portion 72 c between the largerdiameter part 72 a and the smaller diameter part 72 b is pressed bythese preload springs 85 . . . is urged forward to bring the slidingfaces 77 of the stationary valve plate 73 and the moving valve plate 74into close contact with each other.

A steam feed pipe 86 connected to the rear face of the valve body 72communicates with the sliding faces 77 via a first steam passage P1formed within the valve body 72 and a second steam passage P2 formed inthe stationary valve plate 73. Among the casing body 12, the rear cover18 and the rotor 22, there is formed a steam discharge chamber 88 sealedwith a sealing member 87. The steam discharge chamber 88 communicateswith the sliding faces 77 via sixth and seventh steam passages P6 and P7formed within the valve body 72 and a fifth steam passage P5 formed inthe stationary valve plate 73.

As is clear from FIG. 5 when referenced together with FIG. 13 and FIG.16, on the mating faces of the valve body 72 and the stationary valveplate 73, there is provided a sealing member 89 large enough to surroundat the same time the connecting part between the high pressure first andsecond steam passages P1 and P2 and that between the low pressure fifthand sixth steam passages P5 and P6. The sealing member 89 is a so-calledC-type seal having a C-shaped section, surrounding the most part of themating faces of the valve body 72 and the stationary valve plate 73. Theopen portion of its C-shape is open inward in the radial direction. Themating faces are also provided with a smaller sealing member 90surrounding only the connecting part between the low pressure fifth andsixth steam passages P5 and P6. The sealing member 90 also is a C-typeseal, and the open portion of its C shape is open outward in the radialdirection.

Therefore, even if high temperature high pressure steam leaks out fromthe connecting part between the high pressure first and second steampassages P1 and P2 to the mating faces between the valve body 72 and thestationary valve plate 73, the high temperature high pressure steam isblocked by the sealing member 89 covering the whole mating faces,thereby preventing the steam from leaking outward. In this case, as theC-shaped opening of the sealing member 89 comprising a C-type seal isarranged inward in the radial direction, even if the high temperaturehigh pressure steam has a force to leak out from inside to outside inthe radial direction, the steam pressure will expand the sealing member89 to increase its sealing capability, thereby reliably preventing thehigh temperature high pressure steam from leaking out.

Also, as the sealing member 90 surrounds the connecting part between thelow pressure fifth and sixth steam passages P5 and P6 opening to themating faces of the valve body 72 and the stationary valve plate 73, anyhigh temperature high pressure steam leaking out from the connectingpart between the high pressure first and second steam passages P1 and P2to the mating faces is prevented from short-circuiting to the lowpressure fifth and sixth steam passages P5 and P6, and thus wastefuldischarge of high temperature high pressure steam is prevented. In thiscase, as the C-shaped opening of the sealing member 90 comprising aC-type seal is arranged outward in the radial direction, even if thehigh temperature high pressure steam has a force to leak out fromoutside to inside in the radial direction, the steam pressure willexpand the sealing member 90 to increase its sealing capability, therebyreliably preventing the high temperature high pressure steam fromleaking out to the fifth and sixth steam passages P5 and P6.

As the sealing member 89 keeps high pressure on substantially the wholearea of the mating faces, namely the area other than the inside of thesmaller sealing member 90 surrounding the fifth and sixth steam passagesP5 and P6, deformation is prevented by uniformly pressing thecarbon-made stationary valve plate 73 having a relatively low rigidity,and bringing it into close contact with the moving valve plate 74 on itssliding face 77 without any gap, thereby preventing the leakage of thehigh temperature high pressure steam and uneven wear of the slidingfaces 77.

The sliding faces 77 of the stationary valve plate 73 and the movingvalve plate 74 of the rotary valve 71 are not always strictly orthogonalto the axis L, but may be slightly inclined as a consequence ofmachining error or uneven wear. In this state, if the stationary valveplate 73 and the moving valve plate 74 turn relative to each other whilesliding on the sliding faces 77, the valve body 72 supported by the rearcover 18 with the sealing member 82 and 83 therebetween oscillate aroundthe axis L using the smaller diameter part 72 b as a fulcrum within thecompression margin of the sealing member 82 and 83.

In this case, as the valve body 72 is engaged with the rear cover 18 bythe three engaging pins 84 . . . arranged at equal intervals around theaxis L, the valve body 72 can oscillate around the axis L, so that it ispossible to increase the compliance of the sliding faces 77, therebypreventing the leakage of the high temperature high pressure steam andsuppressing further uneven wear of the sliding faces 77. In order tofacilitate the oscillation of the valve body 72, it is preferable tobring the positions of the engaging pins 84 . . . as close as possibleto the axis L. Moreover, as the engaging pins 84 . . . are disposed noton the valve body 72 side but on the rear cover 18 side, an increase inthe inertial mass of the valve body 72 can be minimized to furtherenhance the compliance of the sliding faces 77.

Referring again to FIG. 5, five third steam passages P3 . . . arrangedat equal intervals around the axis L penetrate the moving valve plate74, and both ends of five fourth steam passages P4 . . . formed in therotor 22 so as to surround the axis L communicate with the third steampassages P3 . . . and the expansion chambers 43 . . . , respectively.While the parts opening in the sliding faces 77 of the second steampassages P2 are circular, those opening in the sliding faces 77 of thefifth steam passage P5 are formed in an arcuate shape centering on theaxis L.

Next will be described the operation of the expander E according to theembodiment configured as described above.

High temperature high pressure steam generated by heating water in anevaporator flows from the steam feed pipe 86, and reaches the slidingface 77 of the moving valve plate 74 via the first steam passage P1formed in the valve body 72 of the rotary valve 71 and the second steampassage P2 formed in the stationary valve plate 73 integral with thisvalve body 72. The second steam passage P2 opening in the sliding face77 momentarily communicates for a predetermined air intake period withthe corresponding third steam passage P3 formed in the moving valveplate 74 turning integrally with the rotor 22. The high temperature highpressure steam is supplied from the third steam passage P3 via thefourth steam passage P4 formed in the rotor 22, into the expansionchamber 43 within the cylinder sleeve 41.

Even after the communication between the second steam passage P2 and thethird steam passage P3 is cut off along with the rotation of the rotor22, expansion of the expansion chamber 43 causes the piston 42 fittedinto the cylinder sleeve 41 to be thrust forward from the top deadcenter to the bottom dead center, so that the end portion 61 at thefront end of the piston presses the dimple 31 a in the swash plate 31.As a result, the reaction force which the piston 42 receives from theswash plate 31 gives a rotational torque to the rotor 22. Every time therotor 22 turns a ⅕ round, high temperature high pressure steam issupplied to a newly adjacent expansion chamber 43 to drive the rotor 22for continuous rotation.

While the piston 42 having reached the bottom dead center along with therotation of the rotor 22 is pressed by the swash plate 31 to recedetoward the top dead center, low temperature low pressure steam thrustout of the expansion chamber 43 is discharged, via the fourth steampassage P4 of the rotor 22, the third steam passage P3 of the movingvalve plate 74, the sliding faces 77, the arcuate fifth steam passage P5of the stationary valve plate 73 and the sixth and seventh steampassages P6 and P7 of the valve body 72, into the steam dischargechamber 88, and supplied therefrom to a condenser.

When the oil pump 49 provided on the output shaft 32 is actuated alongwith the rotation of the rotor 22, oil sucked from the oil pan 21 viathe oil pipe 52, the oil passage 95 b of the pump body 95 and the intakeport 53 is discharged from the discharge port 54, and is supplied viathe oil passage 95 c of the pump body 95, the oil passage 32 a of theoutput shaft 32, the annular groove 32 b of the output shaft 32, the oilholes 32 c . . . of the output shaft 32, the annual groove 41 b of thecylinder sleeves 41 and the oil holes 41 c . . . of the cylinder sleeves41 to a space between the smaller diameter part 62 b formed in themiddle portion 62 of the piston 42 and the cylinder sleeves 41. Part ofthe oil held in the smaller diameter part 62 b flows through the spiraloil grooves 62 d . . . formed in the middle portion 62 of the piston 42to lubricate the sliding face in contact with the cylinder sleeve 41,and another part of the oil lubricates the sliding faces of thecompression rings 66 and the oil rings 67 provided on the top portions63 of the piston 42 and of the cylinder sleeve 41.

It is inevitable for water generated by the condensation of part of thesupplied high temperature high pressure steam to infiltrate from theexpansion chambers 43 onto the sliding faces of the cylinder sleeves 41and the pistons 42 to be mixed with oil. Therefore, the conditions oflubrication of the sliding faces are poor, but a sufficient oil film canbe maintained to secure the required lubricating performance bysupplying the required quantity of oil from the oil pump 49 through theinside of the output shaft 32 directly to the sliding faces of thecylinder sleeves 41 and the pistons 42. The size of the oil pump 49 canbe therefore reduced.

The oil scraped off the sliding faces of the cylinder sleeves 41 and thepistons 42 by the oil ring 67 flows from the oil holes 63 c . . . formedin the bottom of the oil ring groove 63 b to the hollow spaces 62 awithin the pistons 42. The hollow spaces 62 a communicate with theinside of the cylinder sleeves 41 via the plurality of oil holes 62 c .. . penetrating the middle portion 62 of each piston 42, and the insideof the cylinder sleeves 41 communicates via the plurality of oil holes41 c . . . with the annual groove 41 b in the outer circumferences ofthe cylinder sleeves 41. Although the circumference of the annual groove41 b is covered by the sleeve supporting flange 34 in the middle of therotor 22, oil within the hollow spaces 62 a in the pistons 42 is urgedoutward in the radial direction by a centrifugal force, and dischargedinto the space 68 within the thermally insulating cover 40 through theoil holes 34 b in the sleeve supporting flange 34, because the oil holes34 b are formed in the sleeve supporting flange 34, and the oil is thenreturned therefrom to the oil pan 21 through the oil holes 40 a . . . inthe thermally insulating cover 40. Since the oil holes 34 b are inpositions deviating farther than the outer end of the sleeve supportingflange 34 in the radial direction toward the axis L, the oil positionedoutward from the oil holes 34 b in the radial direction is held by acentrifugal force in the hollow spaces 62 a of the pistons 42.

As described above, the oil held in the hollow spaces 62 a within thepistons 42 and the oil held in the smaller diameter part 62 b on theouter circumference of the pistons 42 are supplied from the smallerdiameter part 62 b toward the top portion 63 in the expansion stroke inwhich the capacities of the expansion chambers 43 increase, and they aresupplied from the smaller diameter part 62 b toward the end portion 61in the compression stroke in which the capacities of the expansionchambers 43 decrease, thereby reliably lubricating the whole area of thepistons 42 in the axial direction. Moreover, the flow of oil within thehollow space 62 a of the pistons 42 enables the heat of the top portion63 exposed to high temperature high pressure steam to be transmitted tothe low temperature end portion 61, thereby avoiding a local temperaturerise in the pistons 42.

When high temperature high pressure steam is supplied from the fourthsteam passages P4 to the expansion chambers 43, the thermally insulatingspace 65 is formed between the top portion 63 and the middle portion 62of each piston 42 facing the expansion chambers 43, and the thermallyinsulating space 70 is also formed in the rotor head 38 facing theexpansion chambers 43. Therefore, the escape of heat from the expansionchambers 43 to the pistons 42 and the rotor head 38 can be minimized tocontribute to improvement in the performance of the expander E.Furthermore, as the large capacity hollow space 62 a is formed withineach piston 42, not only can the weight of the piston 42 be reduced butalso can the thermal mass of the piston 42 be curtailed for a moreeffective suppression of the escape of heat from the expansion chambers43.

As the metal gasket 36 is disposed between the rear sleeve supportingflange 35 and the rotor head 38 to seal the expansion chambers 43, thedead volume around the seals can be reduced as compared with a case inwhich the expansion chambers 43 are sealed by thick annular sealingmembers, thereby securing a large volume ratio (expansion ratio) for theexpander E and enhancing the thermal efficiency to increase the output.Further, as the cylinder sleeves 41 are configured as separate membersfrom the rotor 22, the material of the cylinder sleeves 41 can beselected in consideration of thermal conductivity, thermal resistance,strength, wear resistance and the like, without being restricted by thematerial of the rotor 22. Furthermore, only the worn or damaged cylindersleeve 41 needs to be replaced, resulting in an improved economy.

Moreover, because the outer circumferential faces of the cylindersleeves 41 are exposed through the two notches 57 and 58 formed in theouter circumferential face of the rotor 22 in the circumferentialdirection, not only can the weight of the rotor 22 be reduced but alsocan the thermal mass of the rotor 22 be curtailed to enhance thermalefficiency. Moreover, by causing the notches 57 and 58 to function asthermally insulating spaces, the escape of heat from the cylindersleeves 41 can be suppressed. Furthermore, as the outer circumference ofthe rotor 22 is covered with the thermally insulating cover 40, theescape of heat from the cylinder sleeves 41 can be suppressed even moreeffectively.

As the rotary valve 71 supplies and discharges steam to and from thegroup of axial piston cylinders 56 via the flat sliding faces 77 betweenthe stationary valve plate 73 and the moving valve plate 74, the leakageof steam can be effectively prevented, because the flat sliding faces 77can be readily machined with high accuracy and permit easier control ofclearances than cylindrical sliding faces do. Moreover, as preset loadsare given to the valve body 72 by the plurality of preload springs 85 .. . to generate surface stresses on the sliding faces 77 of thestationary valve plate 73 and the moving valve plate 74, steam leaksfrom the sliding faces 77 can be suppressed even more effectively.

Further, as the valve body 72 of the rotary valve 71 is made ofstainless steel providing a larger thermal expansion amount, and thestationary valve plate 73 fixed to the valve body 72 is made of carbonor ceramic providing a smaller thermal expansion amount, there is apossibility that the centering between them is displaced due to thedifference in thermal expansion. However, as the fixed ring 79 is fixedto the valve body 72 with the plurality of bolts 80 . . . in a state inwhich the stepped portion 79 a on the inner circumference of the fixedring 79 is pressed in and spigot-fitted onto the outer circumference ofthe stationary valve plate 73 and the stepped portion 79 b on the outercircumference of the fixed ring 79 is spigot-fitted onto the outercircumference of the valve body 72, it is possible to precisely centerthe stationary valve plate 73 relative to the valve body 72 by virtue ofthe aligning effect of spigot fitting, thereby preventing the expander Efrom deteriorating in performance by keeping the supply and discharge ofsteam in time. Moreover, the contact faces of the stationary valve plate73 and the valve body 72 can be uniformly brought into close contactwith each other with the fastening force of the bolts 80 . . . , therebysuppressing steam leakage from those contact faces.

Furthermore, since the rotary valve 71 can be attached to or detachedfrom the casing body 12 by merely removing the rear cover 18 from thecasing body 12, maintenance including repairs, cleaning and replacementcan be significantly facilitated. Also, though the rotary valve 71through which high temperature high pressure steam passes is increasedin temperature, oil can be prevented from being heated by the hightemperature of the rotary valve 71 to deteriorate the lubricatingperformance of the swash plate 31 and the output shaft 32 because theswash plate 31 and the output shaft 32 which require lubrication withoil are arranged on the other side of the rotor 22 than the rotary valve71. The oil also performs the function to prevent overheating by coolingthe rotary valve 71.

When assembling the expander E, it is necessary to adjust the magnitudeof the dead volume between the bottom of the cylinder sleeves 41 (i.e.,the lid member 69 supported by the rotor head 38) and the top of thepistons 42, namely the capacities of the expansion chambers 43 when thepistons 42 are at the top dead center. If the shim 97 interveningbetween the flange 32 d of the output shaft 32 and the inner races ofthe combined angular bearings 23 f and 23 r is thinned, the output shaft32 will move forward (rightward in FIG. 1), resulting in a rightwardshift of the rotor head 38 as well, but the dead volume will decreasebecause the pistons 42 are restricted by the swash plate 31 to be unableto move forward. Conversely, if the shim 97 is thickened, the rotor head38 will move backward (leftward in FIG. 1) together with the outputshaft 32, and accordingly the dead volume will increase. As a result, itis possible to adjust the dead volume as desired by merely replacing theshim 97, and the step otherwise needed for dead volume adjustment can beeliminated to achieve a substantial time saving.

Further, as a single shim 97 having a predetermined thickness issandwiched between the flange 32 d of the output shaft 32 and thecombined angular bearings 23 f and 23 r, to adjust the dead volume onlyby fastening with a single nut 98 the front cover 15 incorporating theangular bearing 30 supporting the swash plate 31 and the combinedangular bearings 23 f and 23 r supporting the rotor 22 and the rotor 22incorporating the pistons 42 . . . , the adjustment procedure issimplified as compared with the conventional adjustment procedure inwhich the thicknesses of two shims, front and rear, are individuallyadjusted. Moreover, since the rotor 22 incorporating the pistons 42 . .. can be kept assembled into the casing body 12 when adjusting the deadvolume, the adjusted dead volume can be confirmed while directlywatching the state of contact between the pistons 42 . . . and the swashplate 31.

When the position of the output shaft 32 relative to the combinedangular bearings 23 f and 23 r is adjusted back and forth by varying thethickness of the shim 97, the position of the rotor head 38 at the rearend of the rotor 22 also shifts back and forth, but there is no problemin adjusting the position of the output shaft 32 because the rotor head38 is slidable in the direction of the axis L relative to the inner raceof the radial bearing 24 disposed between it and the casing body 12.

Then, when the pressure of high temperature high pressure steam suppliedto the expansion chambers 43 urges the pistons 42 in the direction ofbeing thrust out of the cylinder sleeves 41, the pressing force of thepistons 42 presses forward (rightward in FIG. 1) the outer race of thecombined angular bearings 23 f and 23 r via the swash plate 31, theangular bearing 30, the swash plate holder 28 and the front cover 15,and the pressing force of the cylinder sleeves 41 reverse in directionto the suppressing force of the pistons 42 presses backward (leftward inFIG. 1) the inner race of the combined angular bearings 23 f and 23 rvia the rotor head 38 and the output shaft 32. Thus, the load generatedby the high temperature high pressure steam supplied to the expansionchambers 43 is cancelled within the combined angular bearings 23 f and23 r, without being transmitted to the casing body 12.

While the rotor 22 constructed of the output shaft 32, the three sleevesupporting flanges 33, 34 and 35, the rotor head 38 and the thermallyinsulating cover 40 is made of a ferrous material whose thermalexpansion is relatively small, the casing 11 which supports the rotor 22via the combined angular bearings 23 f and 23 r and the radial bearing24 is made of an aluminum-based material whose thermal expansion isrelatively large. As a consequence, there arises a difference in thequantity of thermal expansion in the direction along the axis L betweenthe high and low temperatures of the expander E.

The casing 11 which is greater in thermal expansion than the rotor 22expands more than the rotor 22 and its size in relatively increases inthe direction of the axis L when the temperature is high. Conversely,when the temperature is low, it contracts more and its size relativelydecreases in the direction of the axis L. As the casing 11 and the rotor22 are positioned in the direction of the axis L via the combinedangular bearings 23 f and 23 r, the difference in thermal expansionbetween them is absorbed by the sliding contact of the rotor head 38with the inner race of the radial bearing 24, so that an excessive loadis prevented from acting in the direction of the axis L on the combinedangular bearings 23 f and 23 r, the radial bearing 24 and the rotor 22.This not only contributes to an increase in the durability of thecombined angular bearings 23 f and 23 r and of the radial bearing 24,but also to stabilization in support of the rotor 22, therebyfacilitating its smooth rotation. Moreover, it is possible to preventthe fluctuation in dead volume between the top of the cylinder sleeves41 and the top of the pistons 42 accompanying the change in temperature.

The reason is that, supposing that both ends of the rotor 22 arerestrained by the casing 11 to be immovable in the axial direction, asthe casing 11 tends to contract in the direction of the axis L relativeto the rotor 22 when the temperature is low, the pistons 42 whose top isin contact with the swash plate 31 supported by the swash plate holder28 which is part of the casing 11, are pressed backward, and the rotorhead 38 supported by the casing 11 via the radial bearing 24 is pressedforward, so that the pistons 42 are pressed into the cylinder sleeves 41and the dead volume decreases accordingly. Conversely, when thetemperature is high, as the casing 11 tends to extend in the directionof the axis L relative to the rotor 22, the pistons 42 are drawn outfrom the inside of the cylinder sleeves 41, resulting in an increase indead volume, which in turn invites an increase in the initial volume ofhigh temperature high pressure steam in the normal operating state afterthe warming-up, i.e. a drop in thermal efficiency due to a decrease inthe volume ratio (expansion ratio) of the expander E.

By contrast, in this embodiment of the invention, as the rotor 22 issupported in a floating state in the direction of the axis L relative tothe casing 11, the gaps between the combined angular bearings 23 f and23 r and the radial bearing 24 are prevented from widening and so arethe preloads from decreasing, and the dead volume is prevent fromfluctuating due to temperature change. This enables the volume ratio(expansion ratio) of the expander E to be prevented from fluctuating,thereby achieving a stable performance.

Especially, for the expander E which uses high temperature high pressuresteam as the working medium, the above-described advantage is highlyeffective because the difference is wide between high temperature andlow temperature. Furthermore, although the difference between hightemperature and low temperature is particularly wide in the vicinity ofthe rotary valve 71 to which high temperature high pressure steam issupplied, the difference in thermal expansion between the casing 11 andthe rotor 22 can be absorbed without problem because the rotor head 38can be in sliding contact in the direction of the axis L with the radialbearing 24 arranged closer to the rotary valve 71.

Further, out of the stationary valve plate 73 and the moving valve plate74 of the rotary valve 71, as the stationary valve plate 73 supported bythe casing 11 is urged by the springing force of the preload springs 85. . . toward the moving valve plate 74 supported by the rotor 22, thesealing performance of the sliding faces 77 of the stationary valveplate 73 and the moving valve plate 74 will not be affected even if thepositional relationship between the casing 11 and the rotor 22 in thedirection of the axis L varies along with temperature variations. Notonly that, an excessive load is prevented from acting on the combinedangular bearings 23 f and 23 r and the radial bearing 24, resulting instabilization of the rotational plane of the rotor 22 and accordingly inan improvement in the sealing performance of the sliding faces 77, toreduce the quantity of leaked steam.

Although the preferred embodiment of the present invention has beendescribed above, the invention may be modified in various ways withoutdeviating from the subject matter.

For example, the rotating fluid machine according to the invention isnot limited to the application to the expander E, and is also applicableto a compressor, a hydraulic pump, a hydraulic motor and the like.

Although the expander E in the embodiment is provided with the group ofaxial piston cylinders 56 as the working section, the structure of theworking section is not limited thereto.

1. A rotating fluid machine comprising: a casing; a rotor rotatablysupported by the casing; a working section disposed on the rotor; and arotary valve, provided between the casing and the rotor, for controllingthe supply and discharge of a working medium to and from the workingsection, the rotary valve being constructed by bringing, into contact onsliding faces, a moving valve plate provided on the rotor and astationary valve plate provided on a valve body, and high pressureworking medium passages and low pressure working medium passagespenetrating mating faces of the valve body and the stationary valveplate, wherein the outer circumference of the mating faces is sealedwith an annular first sealing member and the circumference of the lowpressure working medium passages is sealed with an annular secondsealing member inside the first sealing member in the radial direction.2. The rotating fluid machine according to claim 1, wherein the firstand second sealing members are C-shaped seals each having an opening inits section, the opening of the first sealing member is arranged inwardin the radial direction, and the opening of the second sealing member isarranged outward in the radial direction.